Gain scheduling technique for a closed loop slip control system

ABSTRACT

The net engagement pressure of a torque converter clutch having a negative slope capacity vs. slip characteristic is controlled to stably regulate the clutch slip, and thereby isolate engine torque perturbations from the vehicle driveline. The pressure is scheduled with a proportional term responsive to transient slip speed deviations and an integral term responsive to steady state slip speed deviations. Feed forward terms responsive to the throttle changes which force steady state slip speed deviations act to increase the speed of response of the integral term.

This invention relates to a system for controlling the slip of a fluidpressure operated torque converter clutch, and more particularly to apressure scheduling technique which compensates for the nonlinearoperating characteristics of the clutch.

Introduced as an efficiency increasing device, the torque converterclutch is a fluid operated friction device engageable to couple theimpeller (input) and turbine (output) of a fluidic torque converter. Inthe usual application, the clutch is either fully released to permitunrestrained slippage between the impeller and the turbine or fullyengaged to prevent such slippage entirely. An unfortunate aspect of fullconverter clutch engagement is that the engine torque perturbations ortorsionals normally absorbed by the torque converter are passed directlythrough the clutch to the remainder of the vehicle drivetrain and mayproduce annoying pulsations therein if not properly damped. This factoroperates to restrict the usage of the torque converter clutch tospecified vehicle operating conditions for which the annoying effectsare minimized. As a result, the potential efficiency gains afforded byengagement of the torque converter clutch have only been realized over aportion of the range of vehicle operation.

To overcome the disadvantages of a fully engaged torque converterclutch, it has been proposed to operate the clutch in a slipping modewherein a predetermined amount of slippage between the torque converterimpeller and turbine is permitted for regulating the torque capacity ofthe clutch. In any such system, the objective is to isolate enginetorque perturbations in the torque converter while passing steady stateengine torque at a slip rate that provides improved torque converterefficiency. Various systems that control the clutch slippage to achievethe above objectives are disclosed in the U.S. Pat. Nos. to Annis et al.3,730,315 issued May 1, 1973; Cheek 3,752,280 issued Aug. 14, 1973;Chana 3,977,502 issued Aug. 31, 1976; and Malloy 4,181,208 issued Jan.1, 1980, such patents being assigned to the assignee of the presentinvention.

In systems of the above type, it is generally recognized that theefficiency gain due to the control is maximized by regulating the slipspeed as low as possible without fully engaging the clutch. However, thegain, or transfer function, of a typical clutching device tends to benonlinear in nature, and the potential for instability at low slipspeeds is substantial. Characteristically, the fluid pressure requiredto maintain a given level of slippage tends to decrease as the slippagedecreases. As a result, there is a tendency for the control to overshootand completely engage the clutch in response to a condition for whichthe measured slip exceeds the desired slip.

This invention is directed to a closed loop slip control systemincorporating an improved pressure scheduling technique for improvingthe stability of control. The system gain comprises proportional andintegral terms derived as a combined function of the slip command andthe slip error so as to compensate for the characteristically nonlinearclutch coefficient of friction. In addition, the integral term isadjusted by an open loop feed forward technique in response to indiciaof increased torque perturbation, thereby improving the response of thesystem to variations in steady state engine torque, and decreasing theamount of engine torque perturbation transmitted to the drivelinethrough the torque converter clutching device.

IN THE DRAWINGS

FIG. 1 is a schematic and cross-sectional diagram depicting the torqueconverter clutch, certain transmission fluid handling elements, and amicrocomputerbased control unit for carrying out the control functionsof this invention.

FIG. 2 is a graph depicting the torque capacity of the torque converterand clutch as a function of the torque converter slip N for variouslevels of clutch engagement pressure P₁ -P₃.

FIGS. 3a, 3b and 4 are graphs depicting cyclic engine torqueperturbation and the effect of such cyclic variations on converter slip.

FIG. 5 is a control system block diagram illustrating the closed loopslip control system of this invention.

FIGS. 6-7 are graphs depicting the control system proportional andintegral gain as a function of commanded slip and slip error.

FIGS. 8-12 are flow diagrams representative of program instructionsexecuted by the control unit of FIG. 1 for carrying out the controlfunctions of this invention.

Referring now to the drawings, and more particularly to FIG. 1,reference numeral 10 generally designates a portion of an automatictransmission including a conventional fluidic torque converter 12disposed within the transmission housing 14. The impeller 16 of torqueconverter 12 is connected to be rotatably driven by the output shaft 18of engine 20 through the input shell 22, as indicated by the broken line24. The turbine or output member 26 of the torque converter 12 isrotatably driven by the impeller 16 by means of fluid transfertherebetween, and is connected to rotatably drive the torque converteroutput shaft 28 through a splined hub member 30. The stator 32 redirectsthe fluid between the impeller 16 and the turbine 26 and is connectedthrough a one-way device 34 and a splined sleeve shaft 36 to thetransmission housing 14 by fasteners 38.

Also, disposed within the housing 14 and preferably integral with sleeveshaft 36 is a bearing support member 40 in which is disposed a needlebearing 42 for rotatably supporting a drive sprocket 44. The converteroutput shaft 28 is splined in the drive sprocket 44, and a chain 46drivingly connects the drive sprocket 44 to a complementary drivesprocket (not shown) for providing a plurality of distinct speed ratiosbetween the converter output shaft 28 and the transmission output shaft(not shown). A suitable gear set and the controls therefor are describedin detail in the U.S. Pat. No. to Koivunen et al. 4.223 569 issued Sept.23, 1980, and assigned to the assignee of the present invention.

Also disposed within the transmission housing 14 is a torque converterclutch assembly, generally designated by the reference numeral 50. Theclutch assembly 50 comprises a plurality of interleaved clutch plates 52alternately connected to an impeller driven support member 54 or aturbine driven support member 56. A piston axially slidable in the areabetween the turbine 26 and the clutch plates 52 defines an apply chamber60 and a release chamber 62.

Fluid is supplied to or exhausted from the clutch apply chamber 60 viathe fluid line 64; fluid is supplied to or exhausted from the clutchrelease chamber 62 via the fluid line 66 and the converter 12. A checkball mechanism 68 permits a one-way flow of fluid from the releasechamber 62 to the apply chamber 60. When the fluid pressure in the applychamber 60 exceeds that in the release chamber 62, there is a resultantforce which tends to move the piston 58 rightward as viewed in FIG. 1,into engagement with interleaved clutch plates 52. Such force tends toreduce the slippage between the impeller 16 and turbine 26 and whensufficiently great, fully engages the clutch 50 to prevent such slippageentirely. When the fluid pressure in release chamber 62 exceeds that inapply chamber 60, there is a resultant force which tends to move thepiston out of engagement with clutch plates 52. Such force tends topermit increased slippage between impeller 16 and turbine 26, and whensufficiently great fully releases the clutch 50 to permit unrestrainedslippage therebetween. In the released condition, the check ball 68unseats and permits relatively high fluid flow through the converter forcooling purposes.

The control system of this invention operates as described below tocontrol the fluid pressure in the apply and release chambers 60 and 62for controlling the amount of slippage between the impeller 16 and theturbine 26. The remainder of the elements depicted in FIG. 1 are devotedat least in part to such purpose.

The input shell 22 has splined thereto a pump drive shaft 69 whichextends coaxial with and through converter output shaft 28 and drivesprocket 44 and which is mechanically connected as indicated by thebroken line 70 to drive a positive displacement hydraulic pump (P) 72.The pump 72 supplies hydraulic fluid from fluid reservoir 74 to thetorque converter 12, the torque converter clutch control valves, thetransmission control valves, the various clutches and brakes of thetransmission gear set, and the transmission lubrication and coolingsystems. The pump 72 is preferably of a variable displacement design,such as that shown in the Schuster U.S. Pat. No. 4,342,545 issued Aug.3, 1982, and assigned to the assignee of the present invention; however,a fixed displacement pump will perform satisfactorily.

A pressure regulator valve (PRV) 76 is connected to the output of pump72 and serves to regulate the fluid pressure (hereinafter referred to asline pressure) in line 78 by controlling the pump displacement and/or byreturning a controlled portion of the pump output fluid to fluidreservoir 74 via line 79. A second somewhat lower regulated pressure(referred to herein as converter feed pressure) is provided at thepressure regulator output port 105. A pressure regulator valve meetingthe above description is described in detail in the above-referenced toSchuster U.S. Pat. No. 4,342,545.

Reference numeral 80 generally designates a spool valve selectivelyactuable to enable or disable engagement of the torque converter clutch50. The valve 80 comprises a spool 82 having four lands 84, 86, 88 and90 formed thereon. The lands 84, 86 and 88 are of equal area, and theland 90 is of significantly larger area as shown. Line pressure issupplied directly to the valve chamber 92 via the lines 78 and 96 andthrough a restriction 100 to the valve chamber 102 via lines 78 and 104.Converter feed pressure is supplied to the chamber 94 via the line 105.The valve chamber 106 is exhausted to the fluid reservoir 74 through therestriction 107 and the exhaust line 108. The valve chamber 110 isconnected to the apply chamber 60 via fluid line 64, and the valvechamber 112 is connected to the converter 50 and release chamber 62 viafluid line 66.

The valve chamber 114 is connected via line 116 to a solenoid operatedtorque capacity control valve generally designated by the referencenumeral 122. The valve chamber 102 is connected via line 124 to theinput port 126 of a solenoid operated pintle valve generally designatedby the reference numeral 128. As explained below, the pintle valve 128functions to enable or disable engagement of clutch 50, and the torquecapacity control valve functions to regulate the net clutch engagementpressure during operation of the clutch 50.

The pintle valve 128 comprises a pintle 130 axially movable within thevalve bore 132 to selectively connect its input port 126 to exhaust line134 which in turn, directs fluid back to the fluid reservoir 74. Aspring 136 resiliently biases the pintle 130 upward into engagement withthe valve seat 140 as viewed in FIG. 1 to isolate the input port 126from the exhaust line 134, and a solenoid coil 138 is electricallyenergizable via conductor 139 to overcome the resilient force of spring136 and move the pintle 130 downward to connect the input port 126 withthe exhaust line 134.

In FIG. 1, the solenoid coil 138 is depicted in the deenergizedcondition wherein the pintle 130 engages the valve seat 140 to isolateinput port 126 from exhaust line 134. As such, fluid pressure is appliedto the valve chamber 102 of valve 80 via the restriction 100, and thespool 82 thereof is directed upward as shown. In such position,converter feed pressure is supplied from line 105 to the release chamber62 via line 66, and the apply chamber 60 is exhausted to an oil cooler(not shown) via line 64 and exhaust line 109. Fluid from the releasechamber 62 enters the apply chamber 60 through the check ball mechanism68 and is returned to the fluid reservoir 74 via the oil cooler.

When the solenoid coil 138 is energized, pintle 130 moves downward andthe fluid pressure in the valve chamber 102 is exhausted to the fluidreservoir 74 through exhaust line 134. In such case, the line pressurein valve chamber 92 urges the spool 82 downward. In such position, therelease chamber 62 is exhausted through the restriction 107 and linepressure is directed to the apply chamber 60 via the torque capacitycontrol valve 122 and lines 64 and 116. As explained below, the torquecapacity control valve 122 is operated to regulate the pressure in applychamber 62 to control the net engagement pressure of clutch 50.

The torque capacity control valve 122 is a linear solenoid valvecomprising a spool 150 axially movable within the valve bore 152 and apair of unequal area lands 154 and 156 formed thereon. The line pressureconduit 78 is connected to the input port 158 of valve 122, and the line116 from spool valve 80 is connected to the output port 160. Theposition of valve spool 150 is controlled by balancing the differentialarea force with electromagnetic force so as to regulate the pressure inline 116. A spring 166 urges the spool 150 to the left as shown in FIG.1, and a solenoid coil 168 is energizable via conductor 169 to move thevalve spool 150 to the right against the force of spring 166.

In practice, the voltage applied to coil 168 is pulse-width-modulated tocontrol the energization current, and the average current leveldetermines the electromagnetic force to regulate the lineal position ofspool 150 within the valve 122. Relative rightward movement of the spool150 decreases the engagement pressure; relative leftward movement of thespool 150 increases the pressure.

The energization of solenoid coils 138 and 168 is controlled by anelectronic control unit via lines 139 and 169, respectively. The controlis made in response to a number of input signals including a brakesignal (BR) on line 186, an engine throttle signal (TPS) on line 187, atransmission gear signal (GEAR) on line 188, an engine speed signal(N_(e)) on line 189, and a turbine speed signal (N_(t)) on line 190. Thebrake signal may be obtained with a switch mechanism (not shown)responsive to movement of the vehicle brake pedal such that depressionof the brake pedal causes a change in the output state of the brakesignal. The engine throttle signal may be obtained with a suitabletransducer 192, such as a rotary potentiometer (not shown) responsive tothe position of the accelerator pedal or engine throttle 194 forproducing an electrical output signal in accordance therewith. Thetransmission gear signal may be obtained with suitable pressuretransducers (not shown) located with respect to the fluid conductingpassages of the transmission 10 in a manner to determine which gearratio is engaged. The turbine speed and engine speed signals areobtained from speed transducers 196 and 198, respectively.

The speed transducers 196 and 198 may be of the variable reluctance typewhich cooperate with magnetic gear teeth formed on the surface of arotating shaft. Thus, the speed transducer 196 cooperates with the gearteeth 200 of the drive sprocket 44, and the speed transducer 198cooperates with the gear teeth of the engine flywheel or other likerotating member. Alternatively, engine spark firing pulses may be used.

As indicated in FIG. 1, the electronic control unit 180 essentiallycomprises a microcomputer (uC) 202 and an input/output (I/0) device 204,which communicates with microcomputer 202 via an address and control bus206 and a bi-directional data bus 208. A high frequency clock 210supplies microcomputer 202 with a high frequency pulse train forcontrolling the operational timing of the same. The brake, throttle,gear, engine speed and turbine speed signals on lines 186, 187, 188, 189and 190 are applied as inputs to input/output device 204, andinput/output device 204 includes circuitry for converting analog inputsignals to a digital format and for developing suitable control signalson lines 139 and 169 for controlling the energization of solenoid coils138 and 168 in response to duty cycle commands developed bymicrocomputer 202. A flow diagram representative of suitable programinstructions executed by microcomputer 202 in the performance of thecontrol functions of this invention is given in FIGS. 8-12.

FIG. 2 depicts the combined torque capacity of the torque converter andclutch assembly of FIG. 1 as a function of the torque converter slip|N_(e) -N_(t) | for various levels of net clutch engagement pressure.The traces 220-224 depict the combined torque capacity for increasinglevels of net engagement pressure P₁ -P₃. Trace 226 depicts the combinedtorque capacity when the clutch assembly 50 is fully released, andtherefore represents the torque component of the torque converter 12.

At relatively high slip speeds (greater than about 100 RPM in FIG. 2),the torque converter 12 supplies a significant torque capacitycomponent, and the torque capacity vs. slip speed relationship has apositive slope. That is, the torque capacity at a given net engagementpressure increases with increasing slip speed, and decreases withdecreasing slip speed. This relationship describes the operatingcharacteristics of a typical torque converter.

At relatively low slip speeds (less than about 100 RPM in theillustrated embodiment), the clutch 50 supplies the dominant torquecapacity component, and the torque capacity vs. slip speed relationshipmay have a negative slope. That is, the torque capacity at a given netengagement pressure increases with decreasing slip. This relationshipdescribes the operating characteristics of typical friction clutchassemblies.

The significance of the clutch and torque converter operatingcharacteristics described above is graphically illustrated in FIGS.3a-3b. The graph of FIG. 3a depicts engine torque as a function of time.The steady state engine torque is designated T_(avg) and the minimum andmaximum torque excursions are designated T_(min) and T_(max),respectively. The trace 228 in FIG. 3b depicts the combined torquecapacity of the torque converter 12 and clutch assembly 50 as a functionof slip speed for a given net engagement pressure, as in FIG. 2. Thesame scale is used in the abscissa of each graph.

FIGS. 3a-3b illustrate that the steady state engine torque T_(avg) maybe matched by the torque converter 12 and clutch 50 at three differentoperating points: A, B, and C. At operating point A, the average slipspeed N_(A) is about 40 RPM; the clutch 50 is capable of isolatingengine torque perturbations while providing a significant gain in thedrivetrain efficiency. At operating point B, the average slip speedN_(B) is about 170 RPM; the clutch 50 isolates engine torqueperturbations, but provides only a marginal gain in drivetrainefficiency. At point C, the clutch 50 is fully engaged, and transmitsall engine torque perturbations. Obviously, it would be mostadvantageous from an efficiency standpoint to control the clutch tooperating point A.

Notwithstanding the above, stability concerns indicate that it would bemore advantageous to control the clutch to operating point B, where thetorque capacity vs. slip speed relationship has a positive slope. Atoperating point B, increases in slip speed caused by positive inputtorque disturbances such as the positive engine torque excursionsdepicted in FIG. 3a result in an increase in the combined torquecapacity of clutch 50 and torque converter 12. This excursion isdesignated by the point B₁, and is considered to be a stable operatingcondition because the torque capacity of the converter/clutch matchesthat of the engine torque excursion (T_(max)), and tends to reduce theslip speed back to N_(B) when the input torque decreases toward thesteady state level T_(avg).

Similarly, decreases in slip speed caused by negative input torquedisturbances, such as the negative engine torque excursions depicted inFIG. 3a, result in a decrease in the combined torque capacity of clutch50 and torque converter 12. This excursion is designated by the pointB₂, and is also considered to be a stable operating condition becausethe torque capacity of the converter/clutch matches that of the engine(T_(min)) and permits the slip speed to return toward N_(B) when theengine torque increases toward the steady state level T_(avg).

A similar analysis respecting the operating point A reveals it to be aninherently unstable control point. Increases in slip speed caused bypositive input torque disturbances such as the positive engine torqueexcursions depicted in FIG. 3a result in a decrease in the torquecapacity of the converter/clutch. This excursion is designated by thepoint A₁, and corresponds to a slip speed of N_(A1). It is an unstableoperating condition because the decreased torque capacity of the clutch50 permits a further increase in the slip speed. At slip speed N_(A1), atorque capacity of T_(max) is required to match the engine torque, andthe converter/clutch can only generate that much torque capacity(designated by A₁ ') if the net engagement pressure is increased toprovide the capacity vs. slip relationship depicted by the broken trace230. If no torque capacity adjustment is made, the system will quicklyshift to the stable operating point B.

Decreases in slip speed caused by negative input torque disturbancessuch as the negative engine torque excursions depicted in FIG. 3a resultin an increase in the torque capacity of the converter/clutch. Thisexcursion is designated by the point A₂, and corresponds to a slip speedof N_(A2). It is also an unstable operating condition because theincreased torque capacity causes a further reduction in slip speed. Atpoint A₂, a torque capacity of T_(min) is required to match the enginetorque, and the torque capacity of converter/clutch can only be reducedto that value (designated by A₂ ') if the net engagement pressure isdecreased to provide the capacity vs. slip relationship depicted by thebroken trace 232. If no torque capacity adjustment is made, the slipspeed will go to zero, as indicated by point C, fully engaging converterclutch 50.

The converter/clutch operating characteristics described above result ina control dilemma. Efficiency dictates that the slip be controlled inthe negative slope region at less than about 100 RPM, and stabilitydictates that the slip be controlled in the positive slope region atgreater than about 100 RPM.

This invention is directed to a control technique which permits stableslip control in the negative slope region at slip speeds of less thanabout 100 RPM. More particularly, this invention is directed to apressure scheduling technique for controlling the capacity of the clutch50 so that the control behaves as though the capacity vs. sliprelationship had a positive slope. Essentially, the net engagementpressure adjustments are made according to the combination of aproportional gain term G_(p) and an integral gain term G_(i). Theproportional gain term produces pressure changes which account fortransient fluctuations in the slip speed at steady state engine torque,and the integral gain term produces pressure changes which account forchanges in the steady state engine torque.

The proportional gain term G_(p) for a given value of commanded slipspeed is empirically derived as graphically illustrated in FIG. 4. Thecapacity vs. slip speed traces 228-232 of FIG. 3b have been repeated inFIG. 4. For convenience, the trace 232 corresponds to a net engagementpressure P₁, the trace 228 to a net engagement pressure P₂, and thetrace 230 to a net engagement pressure P₃. Essentially, the proportionalgain term G_(p) for a commanded slip of N₁ is graphically represented bythe trace 234. The slope of the trace 234 is opposite in sign and atleast as great as the slope of the traces 228-232 at the commanded slipvalue N₁. Any deviation of slip speed from the commanded value of N₁represents slip error, and determines the net clutch engagement pressurerequired to maintain stable control at the commanded slip value.

Traces such as the trace 234 are defined for various commanded slipvalues (since the slope of traces 228-232 vary with slip speed) and usedto generate look-up table information as graphically depicted in FIG. 6.The table information is referenced to zero, and the information derivedtherefrom indicates the required change in net engagement pressure perRPM of slip error, or in other words, the required gain.

Applying the table information of FIG. 6 to the graph of FIG. 4, assumeas above that the commanded slip speed N₁ corresponds to 40 RPM. If anegative going torque fluctuation causes the slip speed to decrease by10 RPM to N₁ ', the slip speed error E_(s) is -10 RPM, and the netengagement pressure is decreased at a proportional gain of approximately0.7 PSI/RPM slip error. This causes the capacity vs. slip relationshipto assume the level of trace 232 or lower, and tends to drive the slipspeed back to the commanded value N₁. An equal and opposite controlresponse is made when there is a positive slip speed error. However, thegain values for positive slip speed errors--slip speed greater thancommanded--are reduced somewhat as seen in FIG. 6 since relatively largepositive slip error can occur without serious control consequences.

The integral gain term G_(i) shown in FIG. 7 is used to vary the controlpressure as a function of slip error and time. For positive slip error,the gain G_(i) is determined solely as a function of the slip speederror E_(s). For negative slip error, the gain G_(i) is determined as afunction of the percentage of slip error, or (E_(s) /N_(cmd)). As aresult, a given magnitude of negative slip error yields a much higherintegral gain than the same magnitude of positive slip error. Thisasymmetry is needed because a relatively small amount of negative sliperror may result in complete engagement of the clutch 50. Positive sliperror, on the other hand, can be corrected more slowly with relativelylower gain, and substantially no adverse impact on the driveline torque.In either event, however, the authority of the integral gain G_(i) issignificantly greater than that of the proportional gain G_(p) becausevariations in steady state torque which result in the generation ofintegral gain are significantly greater than typical transient torquevariations.

The integral gain term is modified by a feed forward technique toincrease the sensitivity of the pressure control to changes in thesteady state engine torque. Changes in steady state engine torque areanticipated as a function of throttle position and rate of change ofthrottle position. Essentially, the pressure correction occurs sooner intime than it would if the correction were made solely in response toslip error as described above. This is because the throttle movementprecedes the change in engine torque, which precedes the change in slipspeed. In order to prevent the clutch 50 from fully engaging in responseto a negative going steady state engine torque change, the integrationoutput value is reduced in relation to negative rate of change ofthrottle position. The greater the rate of throttle reduction (closure),the more the integrator output value is reduced. To increase the systemgain in response to large positive going changes, and thereby limit theresulting increase in slip, the integral gain term G_(i) is multipliedby a throttle position dependent factor.

FIG. 5 shows a system diagram of a torque converter clutch 50 as setforth in FIG. 1 and a closed loop slip control system therefor accordingto this invention. The elements are depicted in the slip control mode.The clutch 50 and fluid handling elements are schematically depicted inthe area designated by the reference numeral 240. The control systemincludes a slip command generator 244; feedback elements 246 forgenerating an indication corresponding to the actual slip; gainscheduling and feed forward elements 248 for generating a clutchpressure command; an output function generator 250; and a turbine speedroughness detector 252.

As set forth in FIG. 1, the fluid pressure handling elements include apump 72, a pressure regulator valve 76 for generating line pressure inline 78, and a linear fluid control valve 122 for controlling thepressure in line 64. The piston 58 is represented as an area acrosswhich the apply and release chamber pressures act. The line 64 isconnected to the apply chamber 60, the orifice 61 feeds the releasechamber 62, and the release chamber 62 is exhausted through the orifice107 when the solenoid coil 138 of valve 128 is energized for clutchengagement. The torque capacity T_(cap) of the clutch 50 is determinedaccording to the product of the force F acting across the clutch platearea, the coefficient of friction u, and the clutch plate radius RAD.The coefficient of friction u of the clutch plates 52, in turn, isdetermined as a function of the slip (N_(e) -N_(t)) thereacross asindicated by the block 260 and the summing junction 262.

The slip command generator 244 generates a slip speed command N_(cmd) inresponse to engine throttle position TPS (line 187), and turbine speedN_(t) (line 190). Throttle position is indicated along the ordinateaxis, and the various traces correspond to different values of turbinespeed N_(t). In general, the slip command N_(cmd) increases withincreasing throttle position, and decreases with increasing turbinespeed.

The feedback elements 246 include the low pass filters 264 and 266 forpassing the steady state values of measured engine speed N_(e) andturbine speed N_(t), and the summing junction 268 for differencing thefiltered speed values to provide a filtered slip speed indication online 270. The commanded slip N_(cmd) is applied to the summing junction272, where it is combined with the output of turbine roughness detector252 and subtracted from the filtered slip indication (line 270) to forma signal indicative of the slip error E_(s). The turbine roughnessdetector 252 is responsive to an unfiltered measure of turbine speedN_(t) and effectively increases the clutch slip when unacceptableturbine speed variation is sensed.

The proportional and integral gain scheduling portion of the elements248 include proportional and integral function generators 276 and 278,an integrator 280, and a summing junction 282 for generating a pressurecommand P_(cmd) as a function of proportional and integral terms P_(p)and P_(i). The function generators 276 and 278 generate proportional andintegral gain terms G_(p) and G_(i) as described above in reference toFIGS. 6 and 7 as a function of the slip command N_(cmd) and the measuredslip error E_(s). The proportional term P_(p) is determined according tothe product of the proportional gain term G_(p) and the slip errorE_(s). The integral term P_(i) is determined by integrating the productof the integral gain term G_(i) and the slip error E_(s) at block 280.

The feed forward portion of the elements 248 includes the functiongenerators 284 and 286. The function generator 284 serves to reduce theoutput value of integrator 280 in response to the detection of throttlemovement in a negative or closing direction, thereby providingrelatively fast reduction of the clutch pressure when a reduction in theengine torque output is anticipated. The faster response provided by thefeed forward function generator 284 serves to avoid full engagement ofthe clutch 50 when the throttle is suddenly released. The input forfunction generator 284 is provided by subjecting the throttle positionsignal (TPS) to a low pass filter 288, and differencing the filtered andunfiltered throttle signals at summing junction 290.

The function generator 286 serves to increase the integral term P_(i) asthe throttle position is increased from a relatively low setting to arelatively high setting, thereby providing relatively fast increase ofthe clutch pressure when increased engine output torque is anticipated.The faster response provided by the feed forward function generator 286serves to limit engine speed flare (slip) when the throttle is suddenlyincreased.

The output function generator 250 contains the pressure vs. duty cycletransfer functions of the clutch 50 and torque capacity control valve122. It serves to output an energization duty cycle (DC) which, whenapplied to the valve 122, will produce a net clutch engagement pressurethat corresponds with the pressure command P_(cmd). The DC output offunction generator 250 is applied to a pulse-width-modulation (PWM)driver 292, which suitably energized the coil 168 of valve 122.

The control functions described above are carried out by themicrocomputer 202 of FIG. 1 as it executes a set of computer programinstructions stored therein. Such program instructions are functionallydescribed below in reference to the flow diagrams of FIGS. 8-12. FIG. 8represents a main or executive loop, denoting the major tasks performedby the microcomputer 202. The FIGS. 9-12 set forth certain of such majortasks in greater detail.

Referring now more particularly to FIG. 8, the reference numeral 300designates a series of instructions executed at the initiation of eachperiod of vehicle operation for initializing various timers, variablesand registers within the microcomputer. Once the initialization routineis executed, the operations designated by the instruction blocks 302-318are periodically executed in sequence as indicated by the flow lines.

First, the various inputs on lines 186-190 are read and filtered ifnecessary, as indicated by instruction block 302. Instruction block 304indicates the determination of slip speed command N_(cmd) as a functionof the turbine speed N_(t) and the throttle position TPS. Instructionblock 306 indicates the computation of filtered slip speed N_(s) as afunction of the measured turbine and engine speeds N_(t) and N_(e).Instruction block 308 indicates the computation of the turbine roughnessterm TRD, which is detailed in FIG. 9. Instruction block 310 indicatesthe computation of the slip speed error term E_(s). Instruction block312 indicates the computation of the proportional term P_(p), which isdetailed in FIG. 10. Instruction block 314 indicates the computation ofthe integral term P_(i), which is detailed along with the feed forwardfunction in FIGS. 11-12. Instruction block 316 indicates the computationof the pressure command P_(cmd) according to the sum of the proportionaland integral terms P_(p) and P_(i). Finally, as indicated by instructionblock 318, the pressure command P_(cmd) is converted to a duty cycle ofpulse-width-modulation and outputted to the coil 168 of valve 122.

Referring now to the turbine roughness detection routine of FIG. 9, theinstruction block 320 is first executed to compute RPM DELTA, the changein turbine speed since the last execution of the routine. Theinstruction block 321 is then executed to update a long term average,RPM DELTA(AVG) of RPM DELTA. The term RPM DELTA(AVG) may be computed bythe expression:

    RPM DELTA(AVG)=RPM DELTA(OLD)-K1[RPM DELTA(OLD)-RPM DELTA(NEW)]

where K1 is a constant. The instruction blocks 322-323 are then executedto determine a lower limit or threshold of turbine roughness TR LMT as afunction of the turbine speed N_(t), and to adjust TR LMT as a functionof RPM DELTA(AVG). Then TR LMT IS compared with RPM DELTA, as indicatedat decision block 324.

If RPM DELTA exceeds TR LMT, excessive roughness may be present, and theinstruction block 326 is executed to compute a turbine roughnessindication TR according to the expression:

    TR=(RPM DELTA-TR LMT)×G1

where G1 is a fixed gain factor. As explained below, reduction of theoutput TRD of the turbine roughness detection routine is provided byfiltering the roughness indication TR at a variable gain G_(f) accordingto the expression set forth at instruction block 334. However, wheneverthe unfiltered roughness indication TR exceeds the filtered indicationTRD (as determined at decision block 328), instruction block 330 isexecuted to set the output term TRD equal to TR. In addition, the filtergain G_(f) is set at a relatively low value L, and the filter loopcounter, LP CTR, is reset. If TR is less than the output term TRD,instruction block 332 is executed to set the filter gain G_(f) at therelatively low value L, and to reset the loop counter LP CTR. Theinstruction block 334 is then executed to update the output term TRD inaccordance with the expression indicated therein.

If RPM DELTA does not exceed TR LMT, excessive turbine roughness is notpresent, and the instruction block 336 is executed to set the turbineroughness indication TR equal to zero. Then, the blocks designatedgenerally by the reference numeral 338 are executed to set the filtergain G_(f) before the instruction block 334 updates the output term TRDas described above. If the loop counter LP CTR is less than or equal toa first reference REF1 (as determined at decision block 340), theinstruction block 342 is executed to increment LP CTR, and to set thefilter gain G_(f) at a relatively low value L before updating the outputterm TRD. If LP CTR is between REF1 and a second, higher, reference REF2(as determined by decision blocks 340 and 344), the instruction block346 is executed to increment LP CTR, and to set the filter gain G_(f) atan intermediate value M before updating the output term TRD. If LP CTRexceeds REF2, the instruction block 348 is executed to set the gainfilter G_(f) at a relatively high rate H before updating the output termTRD.

Referring now to the computation of the proportional term P_(p) as setforth in the flow diagram of FIG. 10, the decision block 350 is executedto determine the slip error polarity. If the slip error is positive, theproportional gain term G_(p) is determined in accordance withinstruction block 352; if the slip error is negative, the proportionalgain term G_(p) is determined in accordance with instruction block 354.The gain represented by instruction block 352 corresponds to the firstquadrant of FIG. 6; the gain represented by instruction block 354corresponds to the third quadrant of FIG. 6. In either event, theinstruction block 358 is then executed to compute the proportional termP_(p) according to the product (E_(s))(G_(p)).

Referring now to the computation of the integral term P_(i) as set forthin the flow diagram of FIGS. 11-12, the instruction block 364 isexecuted to determine the slip error polarity. If the slip error ispositive, the integral gain term G_(i) is determined in accordance withinstruction block 366 and the instruction blocks 368-370 are executed tocompute the integrator DELTA and to update an intermediate integral termT_(i). If the slip error is negative, the integral gain term G_(i) isdetermined in accordance with instruction block 372, and the instructionblocks 374-376 are executed to compute the integrator DELTA and toupdate the intermediate integral term T_(i). The gain represented byinstruction block 366 corresponds to the first quadrant of FIG. 7; thegain represented by instruction block 372 corresponds to the thirdquadrant of FIG. 7.

Once the intermediate integral term T_(i) is computed, the flow diagramportion depicted in FIG. 12 is executed to complete the calculations forthe integral term P_(i) in accordance with the feed forward indications.First, the instruction block 380 is executed to compute the change inthrottle position TPS DELTA according to the difference between themeasured and filtered throttle values. If TPS DELTA is negative (asdetermined at decision block 382), the instruction blocks 384-386 areexecuted to determine an integrator reduction factor FACTOR and to applyit to the current value of the intermediate integration term T_(i). IfTPS DELTA is positive, the instruction blocks 384-386 are skipped asindicated by the flow diagram line 388. In either event, the instructionblocks 390-392 are then executed to determine throttle position basedgain and to apply such gain to the intermediate integral term T_(i),thereby solving for the signed integral term P_(i).

In a mechanization of this invention substantially as set forth herein,the clutch slip was successfully controlled in the negative slope regionof the capacity vs. slip characteristic as defined above in reference toFIGS. 3 and 4. The pressure scheduling technique of this inventionovercame the natural tendency of the clutch to (1) fully engage inresponse to transient or steady state reductions in engine torque or (2)release in response to transient or steady state increases in enginetorque. The proportional pressure scheduling term has a relatively lowrange of authority but responds quickly to the transient engine torquechanges; the integral pressure scheduling term has a relatively highrange of authority and responds to the steady state engine torquechanges. Both the proportional and integral terms are determined inresponse to deviations in the slip speed of the clutch. The feed forwardterms which are responsive to the throttle changes that force steadystate changes in the slip speed, act to increase the speed of responseof the integral term.

While this invention has been described in reference to the illustratedembodiment, it will be recognized that various modifications theretowill occur to those skilled in the art. Systems or methods of operationincorporating such modifications may fall within the scope of thisinvention, which is defined by the appended claims.

The embodiments of the invention in which an exclusive property orprivilege is claimed are defined as follows:
 1. In a motor vehicleincluding a clutch mechanism connected between input and output membersof a fluid torque converter and adapted to transmit engine outputtorque, the clutching mechanism being characterized in that its torquecapacity is inversely proportional to the slip speed between its inputand output members over at least a portion of the range of such slipspeed, and a clutch actuating mechanism adapted to be supplied withfluid pressure for controlling the torque capacity of said clutchmechanism, a method of operation for controlling the slip speed of theclutch mechanism so as to avoid the transmission of transient engineoutput torque excursions therethrough, the method comprising the stepsof:defining a desired slip signal corresponding to a desired level ofclutch slippage for avoiding the transmission of said transient engineoutput torque excursions through the clutch mechanism, an actual slipsignal corresponding to the actual slip speed of the clutch mechanism,and a slip error signal E_(s) in accordance with the difference betweenthe actual and desired slip signals; supplying fluid pressure to theclutch actuating mechanism in relation to the sum of a first relativelysmall magnitude clutch pressure command, and a second relatively largemagnitude clutch pressure command; generating said first relativelysmall magnitude clutch pressure command as a proportional function ofthe slip error signal for any given value of desired slip, thereby togenerate relatively small magnitude but substantially immediate changesin the pressure supplied to the clutch actuating mechanism when saidtransient engine output torque excursions produce changes in the actualslip; and generating said second relatively large magnitude clutchpressure command as an integral function of the ratio of the slip errorsignal to the desired slip signal, at least when the actual slip signalis less than the desired slip signal, thereby to generate relativelylarge magnitude but integral changes in the pressure supplied to theclutch actuating mechanism when changes in a steady state engine outputtorque produce changes in the acutal slip.
 2. The method set forth inclaim 1, wherein the step of generating the second clutch pressurecommand includes the steps of:generating an integrator gain term G_(i)indicative of a corrective rate of change of clutch torque capacity,such gain term being determined as a function of the slip error signalwhen the actual slip signal is at least as great as the desired slipsignal, and as a function of the ratio of the slip error signal to thedesired slip signal when the actual slip signal is less than the desiredslip signal; forming the second clutch pressure command P_(i)substantially according to the expression:

    P.sub.i =P.sub.i -[G.sub.i *E.sub.s ].


3. The method set forth in claim 2, wherein the motor vehicle engineincludes a movable throttle, the position of which is increased toincrease the engine output torque and decreased to decrease the engineoutput torque, and the method includes the steps of:monitoring thethrottle movement, and generating a first feed forward factor % INT REDin relation to the rate of change of throttle movement when the throttleposition is being decreased; and adjusting the second clutch pressurecommand P_(i) substantially according to the expression:

    P.sub.i =P.sub.i * (1-% INT RED)

whereby the torque capacity of the clutch mechanism is rapidly reducedto avoid full engagement of thereof in response to a sudden decrease inthe throttle position.
 4. The method set forth in claim 2, wherein themotor vehicle engine includes a movable throttle, the position of whichis increased to increase the engine output torque and decreased todecrease the engine output torque, and the method includes the stepsof:monitoring the throttle movement, and generating a second feedforward factor GAIN in relation to the rate of change of throttlemovement at least when the throttle position is being increased; andadjusting the second clutch pressure command P_(i) substantiallyaccording to the expression:

    P.sub.i =P.sub.i * GAIN

whereby the the torque capacity of the clutch mechanism is increased toavoid excessive slippage thereof in response to an increase in thethrottle position.